Centripetal subsonic compressor



July 12, 1955 v. H. PAVLECKA ETAL 2,

I CENTRIPETAL suBs0NIc COMPRESSOR Filed Aug. 12, 1950 e Sheets-Sheet 1 ze- 4 C y 12, 1955 v. H. PAVLECKA EiAL 2,712,895

CENTRIPETAL sussomc COMPRESSOR Filed Aug. 12, 1950 9 Sheets-Sheet 2 y 12, 1955 v. H. PAVLECKA ETAL CENTRIPETAL SUBSONIC COMPRESSOR 9 Sheets-Sheet 3 Filed Aug. 12, 1950 402 4&5

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CENTRIPETAL SUBSONIC COMPRESSOR Filed Aug. 12, 1950 9 Sheets-Sheet s y 2, 1955 v. H. PAVLECKA ETAL 2,712,895

CENTRIPETAL SUBSONIC COMPRESSOR Filed Aug. 12, 1950 9 Sheets-Sheet 6 12, 1955 v. H. PAvLEcKA ET AL CENTRIPETAL SUBSONIC COMPRESSOR Filed Aug. 12, 1950 9 Sheets-Sheet 7 y 12, 1955 v. H. PAVLECKA ET AL 2,712,895

CENTRIPETAL SUBSONIC COMPRESSOR Filed Aug. 12, 1950 9 Sheets-Sheet United .States Patent Ofice 2,712,895 Patented July 12, 1955 CENTRIPETAL SUBSONIC COMPRESSOR Vladimir H. Pavlecka, Pacific Palisades, and Frederick Dallenbach, Inglewood, Calif.

Application August 12, 1950, semi No. 119,028

26 Claims. or. 230-124 This invention relates to novel methods of compressing elastic fluids from low pressures to high pressures, the compression being obtained from the outer periphery of rotary compressors toward their axes of rotation, i. e., the compressions are obtained in radial direction, from the outer periphery toward the center. The disclosed compressors, therefore, are of centripetal type.

This application is a continuation in part of the application Serial No. 557,655, filed October 7, 1944, now Patent No. 2,626,501, entitled Gas Turbine System.

According to one version, the disclosed centripetal compressors comprise a series of concentric and mutually interleaved rings of blades, which have the form of airfoils, arranged in a cascade and concentric relationship with respect to each other. The even numbered rings are integrated by means of a side disc into one rotating unit, while the oddnumbered rings are similarly integrated by means of an identical side disc into the second rotating unit, the two units being rotated in the opposite directions for obtaining high compression ratio per stage, and maximum obtainable overall'compression ratio for the entire compressor. 1

The preferred form of the compressor utilizes two counter-rotating units, while another embodiment of the invention discloses only a single rotating unit, the second unit being stationary.

Overall compression ratios of the order of :1 are ob tainable with the disclosed compressors.

Compression of compressible fluids now is being performed by two types of dynamic compressors known to the prior art, namely axial compressors and centrifugal compressors.

Axial compressors have relatively high efliciency but low compression ratio. The compression ratio of the axial compressors cannot be raised to the desired high ratio because any increase in staging for this purpose produces thickening of the boundary layer, which puts a definite limit to the practicable staging. Centrifugal compressors have relatively high compression ratio but low efiiciency, because of separation of flow in the rotors and stators, this separation being inherent and unavoidable in the centrifugal compressors because of the geometry of the fluid dynamics, and the concomitant thermodynamics, to which these compressors are restricted. Both types of compressors, i. e., axial and centrifugal, are limited in their performance by the unavoidable existence of secondary forces, causing disturbing influences upon the flows within the channels of these compressors, e. g., radial pressure gradient, blade tip leakage, thickening of the boundary layer, the influence from the Coriolis acceleration in the centrifugal flow machines, and the unavoidable three-dimensional flow within the axial flow machines, all tending either to decrease the thermodynamic efliciency or reduce the attainable compression ratio, or both.

The axial flow compressors, besides, are expensive to construct because of twist and changing profile from root to tip of the blades and because a large number of blades is required for attaining even a medium compression ratio. Thus, the axial compressors are excessively large even for medium compression ratios, and the centrifugal compressors are excessively large in diameter and deliver the compressed fluid in the least suitable part of the machine, i. e., at its outermost periphery.

Neither one of these two compressors has the inherently suitable configuration for the compression of elastic fluids, nor are their configurations and basic geometries such as to deliver the products of compression in a suitable form and place for final utilization in the explosion chambers and subsequent turbines or jets with which they are used. Stated more exactly, the products of compression are delivered at a geometrically opposite place than the one that is desired for their proper final fluid dynamic and thermodynamic utilization; According y, all thermodynamic cycles obtainable with the compressors ofthis type are basically fluid dynamically and thermodynamically unsound, because the obtainable geometries lead one completely astray from the paths which are imperative for eflicient thermodynamic and fluid dynamic utilization of the products of compression.

The disclosed centripetal compressors compress elastic fluids from the outermost periphery of its first stage, through a number of rotating radial inflow stages, toward a central chamber situated at the center of rotation. This flow principle determines the functional structure of the centripetal compressors, which arecomposed of a number of rotor stages, each stage having a plurality of airfoils uniformly distributed around the periphery of the stage. The first, outermost rotor stage, is surrounded along its outer periphery by a stationary prerotation stage, which receives air in the radially inward direction,'in oneembodiment of the invention.

It is, therefore, an object of this invention to proyide centripetal compressors having a prerotation stage and a plurality of concentric compression stages.

It is an additional object of this invention to provide centripetal compressors with a prerotation stage producing maximum relative velocity of the fluid and a plurality of compression stages, each stage having a plurality of diffusion channels where high kinetic energy of the fluid, obtained at the exit from the prerotation stage, is converted into a pressure energy through gradual diffusion of the fluid under compression.

It is another object of this invention to provide centripetal compressors having a prerotation stage, said stage having a preacceleration region and an acceleration region, and a plurality of compression stages having an overall maximum compression ratio of the order of 15:1.

Yet another object of this invention is to provide centripetal compressors in which the maximum relative velocity of fluid at the point of entry into the first compression stage is of the order of 0.95 Mach number, in subsonic configuration.

Still an additional object of this invention is to provide centripetal compressors having a plurality of compression stages, each stage having a plurality of flow channels, the maximum angle of turning of these flow channels being of the order of 50, and the percent of maximum diffusion of the order of 3 to 1. 1

Still another object of this invention is to provide centripetal compressors having variable width overall flow channel for introducing volumetric correction into said channel, said correction being determined by the equation of state andcontinuity of the fluid under compression. 1

Yet another object of this invention is to provide centripetal compressors in which each compression stage has the shape of a funnel, with respect to the succeeding stage, in the plane perpendicular to the axis of rotation.

Still another object of this invention is to provide centripetal compressors having a plurality of compression stages, each stage having a plurality of airfoils with the compression stage, and progressively increasing anglesfrom the outer to the inner stage of said .com-

the magnitudes of the respective angles being an inverse function of the radial distance of any given stage c axis of rotation. x n dditional object of this invention is to provide a centripetal compressor having a plurality of compression sta es, stages having constant width channels. I

Other objects and features of the present mventron will appear more fully hereinafter from the following detailed description when considered in connection with the accompanying drawings which disclose .several embodiments of the invention. It is expressly to be understood, however, that the drawings are used here only for the purpose of illustration and are not to be considered as a definition of the limits of the invention, reference for the latterpurpose being bad to theappended claims.

In the drawings, wherein similar reference characters denote similar elements throughout the several views:

Fig. 1 is a vertical axial cross-sectional view of a centtripetal, contra-rotating compressor driven by electric tors; i ig. 2 is a vertical axial cross-section of the upper part of a centripetal contra-rotating compressor having twelve ssion sta F i g 3 is a tr a verse section of a sector of the compressor illustrated in Fig 2;

Fig. 4 is a vertical axial cross-section of a centripetal. single rotation compressor;

Fig. 5 is a transverse section of a sector of the compressor illustrated in Fig. 4;

Fig. 6 is an enlarged transverse section of a prerotation stage, two rotor stages, and one stator stage;

Fig. 7 is an enlarged transverse section of a prerotation stage; i

Fig. 8 is an enlarged transverse section of a single compression stage;

Fig. 9 is a vertical, axial section of the entire flow channel of the compressor illustrated in Figs. 1, 2, and 4;

Fig. 10 is an enlarged transverse section of the prerotation stage and two contra-rotating compression stages;

Fig. 11 is a velocity vector diagram for the stages illustrated in Fig. 10;

Fig. 12 is a complete velocity vector diagram for the compressor illustrated in Figs. 1, 2, and 3;

Fig. 13 is a pressure diagram for centripetal compressors;

Fig. 14 is a volume-pressure diagram for axial, centrifugal'and centripetal compressors;

Fig. 15 is a curve illustrating the acceleration rates obtainable in the prerotation stage;

Fig. 16 is an enlarged cross-sectional view, in section, of the left portion of three compression blades.

Referring to Figs. 1, 2, and 3, they illustrate a twelve stage centripetal compressor. The compressor is mountedon a base frame 15 which supports, by means of four bearing housings, all rotating elements of the compressor plant. Only three bearing housings, 13, 14, and 16 are visible in Fig. l. The rotating elements include two counter-rotating shafts 18 and 20, rotors 22 and 24 of two synchronous motors 26 and 28, four bearings, preferably of the Michell type, only three of which, bearings 30, ,32, and 33 being visible in Fig. 1, two side-thrust bearings 34 and 35, side-discs 36 and 38 of the compressor, and twelve compression stages. 1 through 12, which are illustrated on a larger scale in Fig. 2. Fig. 3 illustrates a transverse cross-sectional view or side-view of the same twelve stages. Each stage is supported from its side disc by means of expansion hoops 40 through 50,'the oddand a corresponding plurality of fluid-turning numbered stages being supported by the side-disc 38, while the even-numbered stages are supported by the side disc 36. The expansion hoops 40 through are integral parts of their respective side discs 38 and 36, the discs and the hoops being centrifugally cast steel castings machined to proper dimensions. The hoops 40 through 50 are tapered toward the center of the compressor to allow relatively free radial expansion of the hoops to follow the radial expansions of rings 51 through 73 which are welded to their respectiveexpansion hoops 40 through 50. The rings 51 through 73 are centrifugally cast steel rings. The inner ends of the rings are used for supporting the compressor blades, such as blades 74 through illustrated in their cross-sectional views in Fig. 3. As illustrated in Fig. 3, the blades are air-foils of a highly cambered profile which will be discussed more in detail in connection with the description of Fig.8.

The airfoil blades are welded to the respective rings and the welded joints are subsequently machined down to smooth finish. The fluid dynamic and thermodynamic characteristics of the compression stages of the compressor will be discussed more in detail in connection with the description of Figs. 7 through 14.

The compressoris surrounded with a prerotational stage consisting of two stationary, peripherally continuous inflow rings 86 and 81 surrounding the first stage of the compressor rotor and a plurality of blades 89, 90, etc., which form acceleration flow channels for radially flowing compressible fluid; these acceleration flow channels also direct the flow of the fluid so that it is in proper vectorial direction with respect to the first rotor stage. The shape of the blades and the behavior of the fluid in the pre-compression stage will be described more fully in connection with the description of Fig. 7.

The prerotation stage is radially attached to frame 15 by means of a plurality of circumferential braces, only four braces 92 through 95 being illustrated .in Fig. l. The entire compressor unit is housed in a hood 96 supported by frame 15, the hood having an opening 97 for admitting the fluid to be compressed into from the prerotation stage.

the compressor, as illustrated by an arrow 98. The

hood opening ,may be provided with two ventilationinlet ducts 99 and 100," which supply proper cooling.

medium for the stators and rotors of the synchronous motors.

The two shafts 18 and 20 are provided with metal linings 101 and 102 and insulating sleevcs 103 and 104 which prevent heat flow from the compressed air into the frame and the rotors of the synchronous motors. The metal linings 101 and 102 constitute the inner, ducts for the compressed fluid which flows in these ducts in the directions indicated by the arrows and 121. The

compressed fluid, upon its emergence from the linings.

101 and 102, reaches stationary diffusing funnels 105, 106, and hoods 107 and 108 which terminate in outlet ducts 109 and 110. Labyrinth seals 111 and 112 are used for making frictionless connection between the rotating elements and the frame of the compressor, ducts 113 and 114 being provided in the frame for conducting the fluid, slowly leaking through the labyrinth seals, into the space confined by the outer cover 115. The cooling air and the air flowing through the labyrinth seals is then conducted outside through ducts or port holes, not illustrated in the drawing. 7

The operation of the compressor is asfollows'. Compressible fluid, which may be air, enters opening .97 and then, at once, is drawn into a pre-compression stage 86878889 where it is imparted sufficient acceleration so as to attain Mach number of the order of 0.7 for the absolute velocity of the fluid at. the exit It is then centripetally compressed, with a substantially constant radial velocity. through all the stages, whereupon it reaches the ducts 101-102 and the outlet ducts 109 and 110. Compresamuse sion ratios of the order of 12:1-l5z1 areobtaintable with the disclosed compressor.

While it is desirable, for optimum operation and ease of construction, to rotate the two compressor rotors at the same revolutions per minute, it is possible to keep one of the compressor rotors stationary. A compressor of this type is disclosed in Figs. 4, 5, and 6. Also, the rotational speeds of the rotors need not be equal; however, when this is the case, the profiles of the airfoils in the two rotors may not be identical but will have to be shaped to conform with the respective rotational speeds of the two rotors. It is obvious from the above, that optimum operating conditions are obtainable only when the two rotational speeds are equal.

Figs. 4, 5, and 6 illustrate a centripetal compressor in which there is only one rotor, the second rotor in this compressor being stationary. In view of the previously given description of the compressor with two rotors, the description of the single rotor compressor may be relatively brief, since the rotors in both cases are identical and the only difference between the two types of compressors resides in the stationary stages of the compressor which now perform only one function, namely to change the direction of flow of the air streams between the successive stages of the rotor so that the angle of incidence of the air velocity vector would conform with the fluid dynamics principles utilized in the disclosed compressors.

Referring to Figs. 4, 5, and 6, and especially Fig. 4, a solid shaft 400 is connected to any prime mover on one side and to a side disc on the other side. The side disc 402, the expansion hoops 403 through 407, and rings 408 through 417 are identical to the same elements in the compressor illustrated in Figs. 1 through 3. The same is true of the compressor blades 500 through 504, etc., Fig. 5, which are mounted between the rings, namely, they are airfoils of a highly camhered profile, which are discussed below and illustrated more in detail in Fig. 8.

The stationary blades 450 through 457 of the com pressor are supported on one side by a frame 424 and rings 460 through 463 and on the other side by rings 425 through 428, these rings, in their shape and function, corresponding to the rotating rings'413 through 417. The compressor is also provided with the prerotation stage 430 supported by frame 424. The compressor is surrounded by a scroll 432 which supplies the compressor with air, if air is the compressible fluid to be compressed. The central chamber 434 of the compressor is connected to an outgoing duct 436, which is connected to any consumer of the compressed air.

Because of the use of scroll 432, the shape of the prerotation stage blades is as illustrated in Fig. 6 at 612 and 614; the blades are straight air foils, and their function is primarily to impart to the incoming air, a velocity having proper vectorial relationship with re-' spect to the air foils 602, 604, 611, and 606 of the first stage of the compressor. The acceleration of the fluid is performed in this case primarily by scroll 432 and only minor degree of acceleration is obtained in the prerotation stage. The stationary blades 600 and 601 have the shape of a constant velocity cascade through which the incoming air is only turned to be in an approximately tangential relationship with respect to the succeeding compression stage of the compressor. The fluid dynamics and thermodynamics of this compressor, in its compression stages, areidentical to those of the contrarotating compressor and, therefore, will not be discussed here. The description of the compression stages of the contrarotating compressor is given below and constitutes a part of the description of the single rotation compression stages.

Fig. 6 illustrates a single stator turning stage 610. The blades 600 and 601 of this stage form a substantially constant velocity channel 603, the function of which is to turn the absolute velocity vector C: so that Fluid dynamics and thermodynamics of the centripetal compressor In any rotary compressors having continuous compression cycle, as differentiated from intermittent cycle (reciprocating compressors), the only way the process of compression can be accomplished is by first accelerating an elastic fluid to as high a velocity as possible to obtain as high a kinetic energy as possible and then converting or transforming this kinetic energy into a potential energy (pressure). The above statement leads one to an immediate conclusion that the higher is the velocity of the accelerated air (if air is the fluid medium) the higher will be the final pressure obtained with the compressor. This being the case, the next logical question is whether there is any practical limit to this velocity. Such absolute limit does exist, and it is within the supersonic velocity range. Supersonic velocity centripetal compressor is disclosed in the pending application of Vladimir H. Pavlecka, filed on March 24, 1951, and having a Serial No. 217,347, where it is stated that the absolute velocity limit isdetermined by the ability of the compressor to cross successfully the boundary between the subsonic and supersonic regions. As also stated in the above application, the optimum performance characteristics are obtainable only at substantially constant speed of rotation. In some applications it is desirable to operate the compressors at variable speed, and when this is the case, then it becomes more advantageous to use subsonic compressors whose characteristics, and especially the airfoils, permit their operation at variable speeds without materially affecting their efficiency. This invention discloses a subsonic compressor.

Whether the compressors are of subsonic or of supersonic type, it is obvious that optimum performance characteristics are obtainable only when applicable laws of fluid dynamics and thermodynamics are observed at every step.

The analysis and a more detailed description of the structure of the compressor will be presented by first discussing the flow channel configurations (Figs. 7, 8. 9, and 15) in the prerotation stage and between two airfoil blades in the compression stage which are capable of satisfying completely the necessary vector relationships between all the entry and exit velocity vectors without any aerodynamic discontinuity throughout the length of the flow channels; the optimum angle of turning of the air fiow .and maximum compressibilities are also discussed in connection with Fig. 8.

This will be followed with the determination of the potential flow field (Fig. 8) to demonstrate that the previously selected parameters satisfy in every respect the laws of fluid dynamics.

Description of the airflow channels will be followed with the description of the velocity vector diagrams (Figs. 10, 11, and 12) throughout the compressor which determine the relationships between the configurations of the mechanical components within the compressor and the laws of fluid dynamics and thermodynamics which must be observed and which control the possible mechanical configurations with the accepted basic geometry of the centripetal compressor. These velocity vector diagrams determine all the air flows and the concomitant fluid dynamic and thermodynamic characteristics of the overall combination. Moreover, these dia- 7 are able to function best within the limits of the vectors. The same principles of analysis will be given for the flow channel form of the compressor in a plane passing through the axis of rotation of the compressortFigs.

9 and 16), the analysis of the flow channels given in connection with Figs. 7 and 8 being'in a plane normal to the axis of rotation.

Prerotation stage of the centripetal compressor tations of fluid dynamics in subsonic region, where Mach number must be considered since it is one of the limiting factors, prevent the assignment to this velocity of a value equal to or in excess of the velocity of sound due to the rounded configuration of the leading edges of the airfoil blades. Therefore, the maximum values of W1 currently attainable have Mach numbers of the order of 0.8 in axial compressors, and 0.95 in one known centrifugal compressor having sharp entry edges (compressor of the De Havilland Ghost jet engine), well known to those skilled in the art. In the disclosed centripetal compressor the relative velocity (W1) Mach numbers'can reach maximum values of the order of 0.9 and 0.95 with rounded entry edges. Therefore, the configuration of the entry channel, which is the channel defined by the surface of the prerotation blade 704 and by the surface of the identical blade 707, must have a pre-acceleration region for gradual acceleration of the air mass crossing the plane defined by the outer periphery of the prerotation stage, and then the acceleration region which enables the fluid to reach the maximum relative velocity having a Mach number of the order of 0.9 to 0.95 at the entry to the first compression stage. The pre-acccleration region is necessary to avoid, in the first half of the entry region, excessive friction losses because of large area exposed to the flow,- and secondly, to avoid flow separation in the intermediate region of the channel where maximum turning of the flow occurs and where. it blends into the accelerating channel proper.

The blades, therefore, assume the shape of sharply curved airfoils which begin with a circular leading edge 708 having .a radius 709 whose center is positioned on a radius 710 of the entire compressor circle 711. The magnitude of this radius 709 is not critical and is primarily dictated, first, by the ability of the cylindrical surface to divide the air flow equally into each channel, and secondly, by the avoidance of flow separation in the process, the latter consideration setting the maximum length that can be assigned to radius 70.

The cylindrical leading edge 708 of the airfoil then blends into two sloping planes 712 and 713 on the opposite sides of the cylindrical leading edge, these planes forming two equal angles, 714 and 715 with radius line 710. The magnitude of these angles is of the order of 8' to 12. 7, then blends into a circle 706 having a radius 716 whose magnitude is not especially critical and is primarily determined by the rate of the pro-acceleration desired in the pre-acceleration region of the flow channel, its minimum value being determined on the other hand by the prevention of harmful encroachment of the pre-acceleration reg'on upon the acceleration region itself and distortion, involving losses, of the blending region between the pre-acceleration and acceleration regions. The inner end of the are produced by radius 716 then forms a tangent connection with a straight line 717, the latter terminating in a semi-circular trailing tip 718, having a radius 724, of the airfoil.

The relative entry velocity, into the @Theatraightlineportion line 72.,the value of this angle beingfrom 2 to S degrees, the optimum anglebeing of the order of 3'. Line 720 parallel to line "721, line 721 being a perpendicular passing thmugh'the. center of line 722. Line 722 is normal to line 728 and also to the convex surface of blade 704. Line 722 represents the minimum width. of the entire flow channel and also the width of the acceleration region. Point 723 represents an intersection of line .722 with the junction point of the lines 717, 720, and semi-circle 718. It is seen from the above that the minimum width of the acceleration region is at the inner end of the concave profile of the airfoil. It is necessary to terminate the airfoil with the finite radius 724 to avoid structural weakness of a thin trailing edge and also to avoid supersonic bending of the outflowing streamand theconcomitant angular deflection of the entire airstream in acounter-clockwisedirection from its intended axis 721. It may be shown that'such deviation, existing in the cascades of the prior art, and currently used in turbines, produces decrease in the overall efficiency of the expansion cascades, of the order of 3%.

The convex surface of the airfoil consists, as stated 2 previously, of cylindrical surface 708, flat surface 713,

. flow-accelerating surface, since it has a pronounced convex curvature, and this has a tendency to produce a vac- The left side of the airfoil, as. viewed in Fig.

uum next to this surface with the result that the air stream accelerates in the direction of the curvature. All that is necessary here is to avoid excessive curvature which would at once produce separationbetween-the stream and the surface, and the resulting turbulence. The magnitude of radius 72$,therefore, is a function of the absolute air velocity. The determination of such surfaces is known to the prior art. Are 728 meets are 727 at point 729, which is the point at which line 722 is perpendicular to a tangent'730' which passes through point 72!. The configuration of the flow channel is such that-tangent 730 is parallel to the axis 721, which is the axis of the how channel. In the priorturbine art, the surface 728 of the air foil follows tangent 730 and the lower portion of the semi-cylindrical tip 718 is tangent to the tangent 730. Stated differently, the .channel wall 728 becomes identical with tangent 730, and therefore is parallel to the direction of the air stream. Because of such configuration, the air stream, not being supported any longer by the trailing surfaces 717 and 717 of the adjacent blade, is deflected downwardly around the tip 718 and such deflection of the air stream carries along with it the entire air stream. Accordingly, the air stream does not follow axis' 721. Such deviation of the air stream produces a wrong angle of incidence on, the following rotor, which is larger. than the calculated optimum angle, which constitutes the second contributing factor to the 3% efficiency loss described previously in connection with the trailing tip 718. Curving ofsurface 728 produces an accelerating surface which counteracts the drooping efiect produced by the trailing tip, and thus restores the direction of the air stream to its intended coincidence with axis 721.

The absolute velocities of the fluid, indicated in Fig.7, are Ca and C1. C0 is the-low entry velocity, while C1 is a high exit velocity whose magnitude is determined by the effectiveness of the first compression stage to create as low a pressure as possible within the air-gap between the two stages. As stated before, the maximum value of C1 is of the order of 0.7 of the local Mach number.

From the description given thus far, it follows that in order to obtainv maximum efiiciency in the prerotation 111 forms an angle 11! with t stage, the stage should have first the preacceleration region which blends into the acceleration region, the flow channel converging gradually through the preacceleration region. Since the acceleration rate is a function of the width of the channel, graphical presentation of the change in width plotted against thelength of the channel will demonstrate, in graphical form, the acceleration rates or the "manner in which the fluid is'accelerated in the prerotationstage channel. Such curve appears in Fig. 15. This curve is obtained as follows: Taking line 722 as a starting point, for example, line 722 is taken as a diameter for circle 732 which is tangent to the adjacent blades at points 723 and 729. Similar circles are inscribed into the flow channel with the centers preferably spaced equidistantly from each other. The line connecting the centers of all the circles is line 721, and this line is the median line of the channel in the plane perpendicular to the axis of rotation. The distance between the circles is then plotted along the ordinate in Fig. 15 and the obtained points on the ordinate are then used to inscribe a plurality of arcs whose radii are equal to the radii of the respective circles illustrated in Fig. 7. Curve 1500 is then drawn as an envelope curve to the arcs drawn previously. The abscissas for the points on this curve are equal to the width of the flow channel divided by two, and the changes in the lengths of these abscissas represent the changes in the accelerations throughout the channel. Examination of curve 1500 discloses that the acceleration changes are very gradual from point 1 -to point 7, are at a maximum from point 7 to point 11, and again rather small from point 11 to point 13, the last portion of the channel thus acting primarily as the flowdirecting channel. Since the length of the ordinate from point 1 to point 13 represents the actual length of the flow channel of theprecompression stage, and since the friction between fluid and the walls of the channel is a function of the velocity, it follows that the shorter is the high velocity region, the smaller will be the loss due to friction. sional changes in the width of the channel to avoid flow separations, and at the same time the dimensional changes were made as abruptly as is possible without producing separations. Thus, the main acceleration region has been restricted to the length equal to the distance between the points 7 and 10 in Fig. 15, which means that relatively high frictional losses are produced in less than one-half of the length of the flow channel.

Compressor rotor stage The accelerated fluid leaves the prerotation stage the absolute velocity C1, which has been discussed in connection with Fig. 7. The same absolute velocity also appears in Fig. 8, which illustrates the transverse crosssectional view of two typical compressor air-foils 800 and 801 taken in a plane perpendicular to the axis of rotation. As indicated in Fig. 8, the rotational velocity of the outer periphery 802 of the compressor stage vectorially is equal to U1. Therefore, the relative velocity of the air stream with respect to the periphery 802 can be represented by a vector W1, the sum of the two vectors U1 and W1 producing the absolute velocity vector C1.

To clarify the representations appearing in Fig. 8, it becomes necessary to digress for a moment and explain the meaning of some of the lines appearing in Fig. 8. The inner periphery of the prerotation stage is indicated at 701, meaning that a radial air-gap indicated by an arrow 803 exists between the prerotation stage and the first compression stage. The radial dimension of the airgap is controlled by the diameter of the semi-circle 718 (trailing end of the prerotation stage blade) which produces a wake behind it. This wake disappears within to diameters of the semi-circle 718, and it is advantageous to make gap 803 sufliciently wide to allow for the complete disappearance of these exit wakes. This streamlines the air stream before it enters the first rotor.

The disclosed channel avoids sudden dimen- Co,=the projection of the absolute velocity C, on the tangent of a circle of radius r,.

. r,=the radius of periphery 802.

The eflect of this is that the absolute velocity C1 in-' creases with the decrease of r,, and angle 0a,, which is the angle between C1 and the tangent, also increases with the decrease of radius r1. Because of relatively smallradial dimension of gap 803, this vectorial change in C1 will be disregarded in this description, although it cannot be disregarded in actual practice.

In Fig. 8 there are indicated the potential flow stream lines 804 through 807 which illustrate the direction of flow only within the lower part of the gap. These lines represent the flow field, or the instantaneous direction of the relative velocity W1 with respect' to the first compression stage. The flow lines 804 through 807 are slightly distorted because of the presence of the convex curvature 830 of the airfoil 801. This distortion extends somewhat into the gap 803 since W1 is less than the speed of sound and therefore the distortion has suflicient time to propagate itself into the gap and in the direction .OpPosite to the direction of flow. This, in turn, will produce the deflection of C1 toward the center of rotation. It is for this reason that the potential flow stream lines assume the form of free vortex having a mean radius The vector velocity diagram is taken at point 808 on the periphery 802 with respect to the mean streamline 805. Therefore, W1 is tangent to streamline 805 at point 808.

It has been stated in the introductory part entitled Fluid dynamics and thermodynamics of the centripetal compressor, that the basic principle of the compression phenomena consists of first accelerating the fluid mass to a high relative velocity and subsequently decelerating it by diffusion, thus converting the kinetic energy into pressure energy. From the above, and previous general,

discussion, it follows that W1 must be as high as possible, and that in the disclosed subsonic compressor this velocity may be of the order of 0.9 to 0.95 of the local Mach number, i. e., Mach number at point 808 in this discussion. This number is a practicable limit for the compressor operating at variable speed in the subsonic region. This Mach number is the highest Mach number that is reached in the compressor and in all subsequent compression stages and interstage gaps the local Mach numher will be progessively smaller.

Proceeding now with discussion of the fluid dynamics of the first compression stage, examination of the fluid potential field discloses that the potential line 809, which is at right angles to the convex surface of the airfoil blade 801 and passing through point 808, is the shortest potential line in the flow field, and thus represents the narrowest portion of the flow channel. Therefore, ac-

cording to Bernoullis principle, the velocity will be- .the greatest and the static pressurewill be the least at this point. Progressing in the centripetal direction, along the mean stream line 805 and into the flow channel of the first stage of the compressor, the lengths of the potential lines 810, 811, 812, and 813 increase progressively, which means that the flow channel ofthe compressor, which begins in the region of the potential line 809 and terminates in the region of the potential line 812, widens progressively from potential line 809 to potential line 812. Such widening of the flow channel potential energy throughout the channel length. The 7 vector diagram of the flow stream at a point 816, which lies at the intersection of'the mean stream line 805 with the inner peripheryline 818, consists of a relative exit velocity W2, peripheral velocity U2, and the absolute velocity C2, which is the sum of W: and Us. Comparison of the vectors at the points 808 and 816 reveals that W1 Wz; U1 U:; and in the illustrated example C1 Ca, all of the above vectors being determined primarily by the dimensions of the flow channel, by; the equations of state for the fluid used,

cFPV=RT (1) and by the equation of continuity and the remaining terms, A,, W,, and are the corresponding values at the potential plane 812.

The pressure and temperature at the potential plane 812 are derived from the equation of state (PV=RT) andfrom the entropy diagrams for the fluid under consideration. The obtained values of P and T are then introduced into the expressions for 'the'density 7 and these in turn are substituted into the continuity equation from which the width of the flow channel at any interstage gapand the velocity vectors are determined.

From the description given thus far, it follows that the flow channel is defined+in the plane normal to the axis of rotation-by the circles 802, 818, and by curvilinear surfaces 820 and 822 of the airfoils 800 and 801. In an ideal centripetal compressor, it would be possible to have maximum relative velocity W1 not at point 808, but at point 824, which is in the vicinity of midpoint of air gap 803. This velocity would then reach point 808 without any change, and-again having an idealized channel in mindthe diffusion and compression would start immediately on the inner side of the potential plane 809 and continue to potential plane 812 and beyond. The airfoils delimiting or defining such channel would have to have zero thickness and could not exert any force on air stream. Therefore, practical airfoils should have suclr fluid dynamic configuration as to have sulficient structural strength and rigidity, and at the same time have a profile which would produce minimum amount of acceleration on the flow stream. This type of airfoil is best approached by the airfoil derived by conformal transformation of a. circle according to Joukovsky, well known to those skilled in the art. Joukovskys airfoils are complex in form and diflicult to fabricate. 9 Accordingly, a better all-way-around air-- possible and, as stated previously, in the theoretical case,

and for fixed speed of rotation, it should be equal to zero, i. e., the leading edge becomes razor sharp. As also stated previously, the subsonic compressor, which is under the discussion here, has its justification in that it is capable of maintaining high efliciency over a wide rotation.

range of rotational speed so long as the leading edge 826 is sufliciently blunt (large radius) to avoid flow separation at its tip. This being the case, the practical leading edge is a compromise, the limit of which nevertheless should be carefully observed, this limit being set by the flow separation.

to the common tangent of both circles at the point where the two circles meet each other. The concave surface 820 is a single circular surface extending from the leading to the trailing tip, its length being indicated by a line 834. Its radius 838 is selected to produce a continuously diverging flow channel. The trailing tip 840 has a small finite radius to give the trailing tip sufficient mechanical rigidity without producing unduly wide wake at exit periphery 818.

The flow channel of the compression stage has been defined thus far in terms of the stream flow lines 804, 805, 806, and 807, potential flow lines 809, 810, 811 and 812, and in terms .of the configuration of the airfoils or blades. It becomes desirable to summarize the full significance of all the terms and analysis of the performance of the compression stage by defining the physical significance of the parameters which spell out the successful and optimum performance of the compression stage. It has been stated before that the shortest potential flow line is line 809 which passes through point 808 determined by the intersection of the mean streamline 805 with the outer periphery 802 of the flow channel. At this point, the relative local velocity W1 is maximum, this maximum being never exceeded anywhere else in the entire compressor. This is so because the width of the flow channel at the potential flow line 809 is at a minimum with respect to any other width in the compression stage. The above means that the potential flow line 809 thusdefines the actual mechanical width of the flow channel at the plane of entry into the flow channel. The same is true of the potential flow line 812 which defines the maximum width of the flow channel at the exit of the compression stage. The above minimum and maximum widths are in a plane perpendicular to the axis of It is the ratio of these widths that determines the degree of diffusion and, therefore, the amount of com pression obtainable in any given compression stage of the compressor. The same limits of the obtainable compres sion may also be defined by the degree of diffusion obtainable in the flow channel, this degree of diffusion being defined by the same ratio.

Examination of Fig. 8 and similar figures (if they were to be drawn in order to demonstrate the practical limits) in terms of maximum obtainable compressions, reveals that this maximum limitis of the order of 2 to l in the subsonic compressor. Itisobvious that it would be use-.

less to include in the above limit a minimum value, since such value would mean making a compressor as inefiicient as possible and still have some output so that it could be called a compressor. parent on the very face of it.

Thus far, the parameters and the performance characteristics of the compressor have been discussed in terms of the lengths of the potential flow lines 809 and 812. The channel obviously has three dimensions and thus defines a total volumetric displacement for any given channel, which is composed of the progressively increasing volumetric increments as one progresses from the entry to the exit of the channel. It goes without saying that the channel must have sides or side restrictions at the sides or ends of the blades.

Subsequent discussion of the channel sides (Figs. 9 and 16) will reveal, however, that the performance characteristics of the channel may best be expressed by the above ratio since it is best if the sides of the channel are given a configuration which makes them not instrument n The main convex surface. of the airfoil preferably is composed of two circular The futility of such effort is ap-- quite fully already, the same conclusions may be arrivedat by following a somewhat difierent analysis, which is.

based on the so-called angle of turning, the term well known in the art relating to axial flow-machines, and which could be borrowed for this use" here. This angle is equal to angle a, which is the angle formed by the potential flow lines 809 and 812. The significance of this angle in the axial flow turbines is that it indicates the degree of energy conversion per stage. In cascade compressors, this angle indicates the. degree of compression obtainable in a single stage, the greater the angle, the greater the compression per stage. Since, in the prior discussion, the amount of obtainable compression ratio has been defined in terms of the ratio of the potential flow lines 812 and 809, it would be only logical to expect that the significance of the value of this angle and of the above ratio is the This is indeed the case, since the geometry of the same. entire compression stage is such that the increase in this angle produces a corresponding increase in the above ratio, i. e., in the width of the flow channel. In the centripetal compressors disclosed here, this angle may have a maximum of the order of 50. W1 being fixed, it is obvious that this angle determines not only the Merit number of the stage, sometimes known also as the pressure coefiicient 11/, which will be defined later (it expresses the degree of compression per stage) but also the number of stages necessary for obtaining the total compression ratio of the entire compressor. Thus, the angle of turning is important not only from the point of view of the merit number, but it is equally important from the point of view of the specific weight" of the compressor, which may be defined as the ratio of the total weight of the compressor to the volumetric flow of the compressed gas per second. Needless to say, this angle will also have a profound eflfect on the cost of the compressor. ()ne of the reasons why the turning angle may have the value which is as large as 50'' is because the airfoils in the disclosed centripetal compressors are turned so as to meet the incoming stream directly with their leading edges and the main surfaces of the airfoils then digress from the natural lines of vortex only to the extent necessary to produce maximum amount of compression without producing flow separations and wakes. Thus, the

compression stage can be turned through a very large angle 61, since the compression configuration of the flow channel turns the free vortex lines even more than in the natural free vortex.

Uniform free vortex approach at the entry into the compression channel around its periphery, and generation of another free vortex at the exit of the same channel, thus aids to obtain compression in the centripetal manner and, therefore, the design of the entire compressor is not With the relative velocity of feet. It is obvious that, considering thenumber of stages, and the inherent complexity of the geometry of the entire compressor, such as twisted blades, it is not diflicult to comprehend the enormous costs and the upkeep difficulties connected with such compressors.

While no direct comparison may be drawn between centripetal compressor and centrifugal compressor in terms of turning angles, since no turning angle'exists in the centrifugal compressors, it may, nevertheless, be stated that staging of the centrifugal compressors represents almost an insurmountable difiiculty from the point of view of fluid dynamics and the delivery of the finally compressed fluid along the axis of rotation. This can be accomplished only by resorting to a 360 turn, which cannot be accomplished without losses.

The previously mentioned turning of the airfoils in the direction of free vortex may be expressed by an angle 43,, between W1 and U1, which angle may be called a scoopmg angle of the compression stage in a centripetal compressor. The significance of this angle resides in the fact that it determines the length of the ditfusion channel; the smaller the angle, the longer the ditfusion channel. This anglemay have a minimum value of the order of 9 in 'the centripetal compressor, while it is of the order of (mean angle) in axial compressors. The length of the flow channel may be defined as the length of the mean stream line 805 between the points 808 and 816. The longer is the flow channel, the larger is the degree of divergence that may be obtained, or, stated again in terms of the ratio of the potential fiow line 812 to line 809, the greater is the length of the fiow channel, the greater is the potential flow line ratio.

beset with the problem of continuously "fighting" the natural laws of flow, which is the case in the existing dynamic compressors, i.e., axial and centrifugal compressors. Because of this inherent disadvantage of the dynamic compressors known to the prior art, the maximum angle of turning in the axial flow compressors is limited to 28, and even then this angle of 28 can be attained only at the root of the blade. section. It means that any subsonic axial compressor must have a large number of-stages, and it is not uncommon in the subsonic axial compressors, having as low a compression ratio as 4.5, to have 11 stages for obtaining this ratio, which makes the overall length of such compressor of the order Profile of the centripetal compressors in a plane passing through the axis of rotation Thus far, the configuration of the compressor prerotation stage and of the compression stages has been discussed with respect to the plane normal to the axis of rotation. The same will be discussed with respect to the plane which passes through the axis of rotation. The configuration of the entire compressor in this plane appears in Fig. l and also, on an enlarged scale, in Figs. 2 and 9, for the contra-rotating compressor, and in Fig. 4 for thecompressor which has one disc, disc 424, stationary and the other disc, disc 402 rotating. For convenient differentiation between the two types of compressors, the contra-rotating compressor will also be called here as the compressor having double-rotation, and the other as a single-rotation compressor. The longitudinal section of the double-rotation compressor will be described first.

In analyzing the compression stage, its functions, and the laws of fluid dynamics and thermodynamics, entirely novel methods have been used here, which are as follows: The configuration of the flow channel from the potential flow line 809 to line 812, Fig. 8, is determined in the manner described in connection with Fig. 8. As stated in connection with the description of Fig. 8, maximum obtainable difiusion is the goal that one strives to achieve in analyzing the entire profile of the flow channel. The obtained compression produces the concomitant changes in temperature, density, specific volume and pressure in the compressed fluid. The above values determine the thermodynamic state of the fluid in accordance with the equation of state (PV=RT) and, since the transverse section (Fig. 8) of each compression stageis dimensioned primarily for obtaining maximum degree of compression rather than its strict compliance with the equation of state,

it becomes necessary to adjust the overall volumetric characteristics of the flow channel in the planes perpendicular to the axis of rotation, i. e., by adjusting the shapes of the side rings 51 through 73, Fig. 2, and the axial dimensions between the rings in each stage. Thus, this volumetric adjustment is produced without unnecessarily burdening the transverse section with such correction, which wouldbe in direct opposition to what one then on.

strives to accomplish in the flow channel. Stated more succintly, there is a necessity of obtaining as large a ratio of 812 divided by 809 as possible to obtain maximum diffusion; yet, the volumetric correction would call for the decrease in this ratio, which obviously either would reduce or completely stop the diffusion, and thus stop the entire cycle of compression. From what has been stated about the vectorial velocities of the compressible fluid in the transverse plane, i. e., the plane illustrated in Fig. 8, it follows that in the centripetal compressors the compression is accomplished by varying the magnitudes of these vectors in this transverse plane only, and this flow does not have any operating component in the axial plane. Thus, the entire compression cycle is based upon the variation of the velocity components in a single plane,

i. e., say, X-Y plane, which is the plane illustrated in Fig.

8. This type of compression cycle may be called the twodimensional compression method as differentiated from the three-dimensional compression method currently used in axial and centrifugal compressors. The above follows at once from the very geometry of the compressors, the centripetal compressor having neither the blade twist nor the change in the airfoil profile, while the centrifugal and axial compressors have both. Thus, in any two-dimensional compressor the volumetric correction cannot be made in the transverse plane (Fig. 8) and the width of the flow channel cannot be made uniform since this at once would destroy the obtained compression by removing the necessary side-support for the compressed fluid.

In view of the above, all volumetric corrections must be introduced in the longitudinal plane. These corrections are illustrated quite clearly for the entire channel in Fig. l, where the width of the flow channel begins to narrow even within the prerotation stage. It is to be noted here that narrowing of the prerotation stage is done tomake its volume conform with the accelerated velocity rather than for obtaining any compression since it is obvious that in this stage the exit pressure is always less than ambient. From then on, however, narrowing of the width of the channel up to its mid-portion, in the volumetric correction.

illustrated. example up to and including the seventh stage 7 (see Figs. 1 and 9) is done because of the compression of the fluid and the necessity of continuously adjusting the volume of the channel to the continuously changing state of the fluid as defined by the state equation and the equation of continuity. Beyond stage '7, the channel begins to widen until it finally reaches the central chamber. The channel must be made wider as one approaches the axis of rotation because the volumetric displacement of the last stages diminishes with such progressive rapidity with. the decrease of their radii that it becomes necessary to counteract this accelerated decrease in volume by widening the channel width. The shape of the resulting curve in the upper part of the channelapproaches a parabola, and it then becomes first almosthyperbolic and then strictly hyperbolic where it makes a turn at the entry of the channel into the hollow part of the. shaft.

Beginning with the description of the profile of the prerotation stage, this profile from point 900 to point 901 is controlled primarily by the profile of the compression stages, since profile 900-901 represents in the main a smooth continuation of the main compression flow channel. The main overall dimension of the first compression stage 1, namely its radial dimension 902 and its axial width 904 at the point of entry are determined by the desired capacity of the compressor, and from then on the profile of the side walls of the entire compression channel from line 905 to line 906 (see stage #12) is determinedby the previously mentioned state and continuity equations, the radial inflow velocity, and decrease in the available ;volume as one progresses in the radial direction.

This profile is represented by a dotted line 908, and, as mentioned previously, the width of the channel diminishes up to stage 7, and then widens from The converging portion of the profile from point 900 to approximately a point 910, on the convergrngportion of the curve, approaches a parabola. The remaining diverging portion of the curve approaches a hyperbola from 910 to 912, and from point 912 to point;

914 it is a hyperbola. The radial component 'Cr, of the exit velocity Ce from the last stage, is relatively small as compared to Co. In order to make the diameter 916 of the ducts in the hollow shafts 18 and 20, Fig. 1, as

small as possible, it becomes desirable to make Co, the axial velocity, as high as possibIe without reaching the absolute maximum limit of flow in a constant diameterchannel, which corresponds to Mach number 1. Mach number 0.8 to 0.9 isa feasible practical limit. The actual shape of the volumetric correction curve, to a large extent, is a function of the number of stages used in the compressor, any increase in the number of stages flattening out the curve, and vice versa. The curve also flattens out with the decrease of the flow capacity of the compressor since the axial dimension of the compressor becomes so small that only small volumetric correction is required, but it is never equal to zero in a practical compressor. 'It becomes equal to zero only when the r capacity of the compressor becomes equal to zero, which is only a logical theoretical conclusion. The dotted line 908 is a theoretically correct boundary defining the In practice, this line must be defined by the inner surfaces of the hoop rings51 through 62 on one side, illustrated in Fig. 9 (see also Fig. 16), and r by the hoop rings 63 through 73 on the other side, which are illustrated in Fig. 2. Since any surface introduces friction, it becomes necessary to make proper allowances for the boundary layer which is induced by the frictional flow of the fluid with respect to the walls of the channels,

line 908 representing the boundary of the channel with the viscosity of the fluid being equal to zero. From the above, it follows that the higher is the viscosity coeficient, the larger is the correction which mustbe introduced for obtaining the desired effective width of the channel. This correction is denoted by a dashed line 918, Figs. 9 and 16, the spacing between line 908 and 918 being equal to the local thickness of the friction boundary layer at any given point. This spacing is not constant, but increases from point 901 to point 910 and decreases to point 914. The increase from point 901 to point 910 is due to the constantly increasing pressure and a corresponding increase in friction between the fluid and the side walls of the channel, this increase in friction producing an increase in the thickness of the boundary layer by producing centrifugal acceleration of the fluid at the point of contact between the fluid and the rotating side wall. This centrifugal acceleration is greater than the opposing centripetal acceleration because-contour 908 is of decelerating type up to point 910. From then on there is a decrease in the thickness of the boundary layer because of the accelerating nature of the profile of curve of the compression stages gets shorter. This progressive decrease in the thickness of the boundary, layer continues 0 even into the region of the hollow duct 916 (its diameter) because of the increase in the absolute velocity 0.,

which produces a marked increase in the Reynolds number (Re) far into the turbulent region of flow with correspondingly low coemcients of friction.

The above described phenomena of progressively decreasing thickness of the boundary layer beyond the region of point 910 toward the axis of rotation is not a mere matter of greater efllciency, although it definitely contributes to the higher efliciency. It affects the very fluid dynamics of the compressor, and therefore must be considered in anriving at the operable compression cycle. Thus, in axial and centrifugal compressors, where the boundary layer becomes progressively thicker in the succeeding stages of the axial compressor and with the increasing radius of the centrifugal compressor, it is a matter of established experimental incontrovertible fact that these boundary layers had rendered some of the compressors completely inoperative. The axial and the centrifugal compressors which are operative solvethis problem by means of first very poor but inescapable compromises, i. e., an increase inthe width of the channel to the extent which materially impairs the efi'iciency of the compressor (even then the velocity diagrams are completely out of balance with the angles of the flow channels) and second, by having very low compression ratios, the maximum compression ratios obtainable being of the order 4.5 as compared to 8 to 15 reached with the compressor illustrated in Fig. 2 when the velocity vector relationship illustrated in Figs. 10 and 11 is used.

The contour line 918 does not represent the actual contour of the walls of the rings 51 through 62 since the actual wall surfaces incorporate additional corrections for reducing the thickness of the boundary layer. The above is accomplished by tilting the surface slopes in the counterclockwise direction as viewed in Figs. 9 and 16. Taking tween the boundary contours 908 and 918, is made narrower at the inner part, i. e., next to the inner edge 926 of the hoop ring 57, and wider at the outer part next to the outer edge 924 of the same ring. Thus the narrowed part represents the region of maximum radial velocity in the boundary layer, which will reduce its thickness, and the wider region acts asa funnel with respect to the inflowing boundary layer from the preceding acceleration region of the preceding stage in the plane perpendicular to the axis of rotation. Y

It should be noted here that angle is not constant for all stages but varies from stage to stage, being least for the outer and inner stages, and maximum in the intermediatestage. This follows from the slope of curve 918 with respect to the central axis 928 of the entire compression channel.

The funnel type of construction of the side walls has an additional'purely accidental advantage, which is not in the realm of fluid dynamics, but'in the realm of ease of construction. If the channel surfaces of the hoop rings had to follow strictly contour line 918, the mechanical tolerances would have to be most exacting for realizing perfect alignment of all the surfaces in the axial direction, i. e in the direction of the axis of rotation 930. Because of the funnel type of construction, the mechanical tolerances for obtaining workable axial alignment of all stages becomes much more liberal since the axial dimension 932, which represents the amount of funnel overlap, is not a very critical dimension.

Velocity vector diagrams of the compressor stages It has been stated in connection with Figs. 7 and 8 that the compressible fluid enters the prerotation stage with the velocity Co and leaves this stage with the velocity C1, both of these velocities being absolute velocities. The fluid then enters the first stage with the absolute velocity C1 and, because of the rotation of the first stage with the peripheral velocity U1, the actual velocity of the fluid with respect to the flow channel of the first stage becomes the relative velocity W1, which is the velocity with which the fluid enters the first flow channel. This is represented by the vector diagrams in Figs. 8, 11, and 12.

Ill)

. In order to simplify the vectorial presentations of all the velocities in all compressor stages, these velocity diagrams have been taken along a radial line 1001 illustrated in Fig. 10. This enables one to make all peripheral velocities Us, parallel to each other and position all vector diagrams along one radial line 1001, which also appears in Figs. 11 and 12. Such presentation does not constitute an approximation, but corresponds to a physical analogy if the airfoils were made infinitely thin and infinite in number. Stated differently, the presence of the airfoils is disregarded without disregarding all the physical effects that they produce.

The first vector diagram in Fig. 11 thus has been taken at a point 1002. The angles formed by the vectors with respect to the tangent lines 1000 and 1020 are A61 and 451 respectively, the first being the angle of the absolute velocity C1, and the second being the angle of the relative velocity W1. These angles were discussed in connection with the description of Figs. 7 and 8. The velocity diagram at the point of exit 1004 from the first compressor stage is represented by the vectors We, U2, and C2, and the velocity diagram for a point 1006 on the periphery of the second stage is represented by the vectors Ca, We, and Us, the absolute velocity C3 being substantially parallel and equal to the absolute velocity C2. The corresponding angles angle angle (:2, angle 53, and angle as indicate the angular changes in the velocity vectors with respect to the tangent reference line. In terms of actual physical angles, the progressive increase in the angles as and Bs with decrease in radius finds its counterpart in the progressively increasing pitch angle c, Fig. 10, which is an angle between the airfoil chord line 1010 and a tangent line at the point of the trailing edge of an airfoil. This angle roughly defines the slope of the mean flow line; in the disclosed compressor angle e is inversely proportional to the respective radii of the compression stages.

The process of the vectorial presentations of the velocities and their angular relationships with respect to the radial line 1001 has been carried all the way through all the stages of the compressor illustrated in Figs. 1 and 2, with the final absolute exit velocity appearing as vector C24. The complete symmetry of all vector diagrams with respect to line 1001 is apparent even from casual examination of Fig. 12, all vectors diminishing linearly withthe decreasing radius of the compression stages, as indicated by the straight lines 1201, 1202, 1203, and 1204, all of these lines converging to a point 1205 which lies on the axis of rotation 1206. Line 1201 makes an angle A with line 1001 and line 1204 makes an angle :2 with line 1001, the two angles being equal to each other. The same is true of angle q and angle I, which are also equal to each other. The significance of the above symmetries in terms of the performance of the respective stages resides in the fact that all stages carry proportional loads, the absolute values of which are a function'of the square of the radius of each stage. The above relationship follows from the application of the fundamental equation for an increment in mechanical head, derived from the Euler's equation for an increment in mechanical heads due to change of velocities in fluids. The general equation reads:

AL=% c.,U.-c.,U, s)

where AL is the mechanical-head in lbs.

g is acceleration due to gravity, feet/sec.

Cu, is the absolute entry velocity vector projected upon the tangential direction, feet/sec.

Cu, is the absolute exit velocity vector projected upon the tangential direction, feet/sec.

U1 is the peripheral-velocity at the entry to the stage,

feet/sec. I

19 U2 is the peripheral velocity at the exit of the stage,

feet/sec.

It may be shown by differentiating Equation 3 for determining the maximum value of AL, that this maximum value is obtained when Ce,Cu,= Us

Actual measurement of these vectors in Figs. 11 and 12 demonstrates that the vector diagrams do satisfy this equation.

The fact that all stages carry proportional loads, the absolute values of which are a function of the' square of the radius of each stage, follows from Equation 3 which states that the mechanical head, AL, is a function of the square of the radius since Cu and U are both functions of the radius.

From the relative magnitudes of the vectors in the first stage and the last stage, one may gain an impression that little is accomplished by the last stages of the compressor. Although AL is relatively small, the work done by these stages is fully in conformance and inproportion to their size so that, relatively speaking, they are as fully loaded as the first stages.

The vector diagrams in Fig. 12 disclose the fact that the projection of the absolute velocities C's on the radial line 1001, identified as the radial velocity Cl in Fig. 11, is substantially constant in all stages, which represents the optimum operating condition for all stages, primarily from the fluid dynamics point of view.

It should be pointed out here for the sake of accuracy that the head equation for the centripetal compressor has one additional term, if it is compared with Equation 3. For the centripetal compressor, the head equation, derived from the Eulers Equation 3, is as follows:

It has been stated in connection with the discussion of the performance of a typical compression stage that maximum attainable compression ratio in the first compression stage is of the order of 2 to l, and that compression ratios of the order of 12 to l are attainable with the twelve stage contra-rotating subsonic compressor.

.One of the main factors contributing to the attainment of such high overall compression ratio resides in the fact that there is a very definite geometric, as well as fluid dynamic, relationship between the prerotation stage and the first compression stage, this relationship dictating the. performance and the merit numbers that must be assigned to the remaining stages in order to carry out what has been' started in the prerotation 'stage and the first compression stage to its logical conclusion. It is for this reason that the vector diagrams illustrated in Fig. 12 assume that perfect symmetry which is illustrated with the striking clarity in Fig. 12, i. e., note the lines 1201, 1202, 1 203, and 1204, and the fact that angle A=angle S2, and angle I =angle I.

The above mentioned geometric relationship resides in the fact that the flow channel of the prerotation stage has a sharp bend. to the right in Fig. 17, as viewed in the direction of flow. This bend, as stated previously, is defined by the radius 716 on the concave portion and by radius 725 on the convex portion of the blade.

This is the beginning of the acceleration channel proper, and from then on the mean flow line 721 assumes the angle -Whih determines the angle of the absolute 6 velocity C; with respect to the input portion of the compression channel in the first compression stage; As illustrated. very clearly in Fig. 10, this velocity, together with the rotational velocity U1 determine the vectorial position of the relative velocity W1 with respect to the mean flow line 805 in Fig. 8, W1 being parallel to line 805 at the point of entry. It means that the compressible fluid enters the first stage with the maximum velocity attained anywhere in the compressor and such entry is directly in line or in the direction of the open flow channel at the point of entry 808 into the first stage. Thus, the flow channel of the first stage becomes fully rammed in with the compressible fluid, and this ramming is then continued throughout the compressor. Maximum kinetic energy is used for supercharging the first compression stage, and the process is continued to the last stage with the diminishing kinetic energies but with the constantly increasing densities and pressures of the fluid.

The above is achieved by tilting the leading edges of all airfoils in the first stage in the direction of rotation and making the centers of. the leading edges point into the acceleration channel of the prerotation stage. The above geometry is defined in terms of the first com-. pression stage looking into the acceleration channel of the prerotation stage. The same may be also defined in terms of the trailing edges of the prerotation stage, which are sharply turned in the direction opposite to the direction of rotation of the first compression stage. This geometry produces the all important supercharging of the compressor with the compressible fluid. The prerotation stage obviously plays an important role in this process since without this stage it is impossible to obtain the all important kinetic energies and the local Mach number of the order of .9 to -.95. The only theoretical substitute for the prerotation stage-would be such high peripheral velocity of the compression stages which would exceed all possible mechanical stress limits of all known materials.

At the conclusion of the descriptions of the physical configuration of the disclosed compressors, it is fitting' to mention, if only parenthetically, that attempts have been made by others to produce centripetal compressors. United States Patents 2,391,779 to Griflith, 2,357,778 to L. W. Beaven, 2,471,892 to N. C. Price, 1,644,565 to P. L. Crowe, and 1,748,979 to Ganderson purport to disclose centripetal compressors. None disclose acceleration channel in prerotation wstage, volumetric correction and properly coordinated diffusion channels in the compression stages. .It is obvious, in the light of the most exhaustive discussion of the laws of -fluid dynamics and thermodynamics applicable to the :centripetal compressors given here, that all. of the above features are'indispensible to produce an operative centripetal compressor.

Appendii:

7 where it is the non-dimensional head coeflicient g is acceleration of gravity, I feet 4 sec.

stage where 'he terms have the definitions given previously in the specification.

The compression head A H can be expressed in terms of temperatures as follows:

AAH=C',,(T T,) (8) mean where l B. t. u. A= mechanical equivalent of heat Erma.

T =temperature of gas at the inlet to the stage, R.; T =temperature of gas at the outlet from the stage, R.; C',,=mean specific heat at constant pressure between mean T1 and As stated previously, for optimum performance of the compressor Cr=constant and Cu,--Cu =U,-

When these two relations are satisfied, the compression head produced by a single rotating stage may be expressed by UtCi S :11

2 9 AH 29 g and the head coeflicient, b, becomes where A is or absolute velocity, 0; divided by peripheral U velocity, U, for the compression stage under consideration.

5 is g U1 and U, having been defined before.

The isentropic head coefiicient, b, is:

2 AH 0 a 73 where 1; is the efiiciency of the compression stage.

AH is the isentropic compression head in feet.

this head difference designating the amount of energy converted into heat by friction while raising the pressure from 121 to p2.

Designating by the velocity ratio W: and substituting this into Equation 13, the latter becomes:

22 and, if

then, from Equations 10, 11 and 14, one obtains t 1 n= p From this equation it follows that the higher the head coefficient and the higher the peripheral velocity, the higher will be the efliciency of the compressor. Thus, for a centripetal compressor, the most important stages have the highest eificiency.

Fig. 13 illustrates a curve of A111, and an increment of head coefiicient with respect to :11. The curve illustrates that with the increase of the depth of the compression stage in the radial direction, the increment of the head coeflicient, Arp, becomes greater in the centripetal compressors, and that any increase in the depth of the compression stage will produce higher efliciency. The increase of the increment A 0 is caused by the increase of the pressure head in the centripetal compressors with respect to the centripetal head.

The dotted line in Fig. 13 connects the points of zero slope on all curves having Ar -1.0 and A l.0,

it: or' 1 '2 Therefore, the abscissa of Fig. 13 corresponds to the case of all axial compressors, while all curves having x 1 correspond to centripetal compressors.

The symbol defines the radial stage depth in terms of the ratio of the outer and inner radii respectively of a stage. The abscissa has the value of A=1.0 and therefore represents any axial compressor. Any number of curves may-be drawn based on values of X=1.0, each curve defining a special radial depth of a centripetal stage. Of these curves only one is shown, for A=l.l0, a value most representative of centripetal compressors. The dotted line, almost vertical to the abscissa, connects the points of zero slope of all the A curves, as indicated onthe ).=l.10 curve.

Volume-pressure characteristic of a centripetal compressor One of the most revealing characteristic curves of any dynamic compressor is the volume-pressure curve which demonstrates the change in pressure produced by the compressor with change in the volumetric flow of fluid through the compressor at some selected speed of rotation, the curve being obtained by throttling the compressor on the output side. Typical curves for three types of compressors are illustrated in Fig. 14. Examina tion of these curves-reveals that all compressors start with some positive pressure at zero volumetric flow, then rise to a maximum pressure and then drop rapidly toward zero pressure at maximum volumetric flow, the latter point being of purely theoretical interest. Practical operation of the centrifugal and axial compressors is possible only in the regions indicated in Fig. 14 which region has a negative slope. Operation of these compressors on the positive slope side of the curvesis' impossible because of flow separations taking place in that region and consequent violent surging of the machines. The operating region of the centripetal compressor is on the positive slope side of its curve, which is very desirmore air with increasing demand for more air from the turbine, and the tremendous problem of matching a tur- 

